DYNAMICALLY TUNED SHROUD FOR ATTENUATING GUN BARREL VIBRATION

Transcription

1 AD TECHNICAL REPORT ARCCB-TR DYNAMICALLY TUNED SHROUD FOR ATTENUATING GUN BARREL VIBRATION A. G. LITTLEFIELD E. L. KATHE R. DUROCHER AUGUST 2002 TACOM-AROEO US ARMY ARMAMENT RESEARCH, DEVELOPMENT AND ENGINEERING CENTER Close Combat Armaments Center Benet Laboratories Watervliet, NY APPROVED FOR PUBLIC RELEASE; DISTRIBUTION UNLIMITED

2 DISCLAIMER The findings in this report are not to be construed as an official Department of the Army position unless so designated by other authorized documents. The use of trade name(s) and/or manufacturer(s) does not constitute an official endorsement or approval. DESTRUCTION NOTICE For classified documents, follow the procedures in DoD M, Industrial Security Manual, Section 11-19, or DoD R, Information Security Program Regulation, Chapter IX. For unclassified, limited documents, destroy by any method that will prevent disclosure of contents or reconstruction of the document. For unclassified, unlimited documents, destroy when the report is no longer needed. Do not return it to the originator.

3 REPORT DOCUMENTATION PAGE Form Approved OMB No Public reporting burden for this collection of information is estimated to average 1 hour per response, including the time for reviewing instructions, searching existing data sources, gathering and maintaining the data needed, and completing and reviewing the collection of information. Send comments regarding this burden estimate or any other aspect of this collection of information, including suggestions for reducing this burden, to Washington Headquarters Services, Directorate for Information Operations and Reports, 1215 Jefferson Davis Highway, Suite 1204, Arlington, VA , and to the Office of Management and Budget, Paperwork Reduction Project ( ), Washington, DC AGENCY USE ONLY (Leave Blank) TITLE AND SUBTITLE DYNAMICALLY TUNED SHROUD FOR ATTENUATING GUN BARREL VIBRATION 2. REPORT DATE August REPORT TYPE AND DATES COVERED Final 5. FUNDING NUMBERS AMCMS No H180.0 PRON No. 4A1C1FYA1ABJ AUTHORS A.G. Littlefield, E.L. Kathe, and R. Durocher 7. PERFORMING ORGANIZATION NAME(S) AND ADDRESS(ES) U.S. Army ARDEC Benet Laboratories, AMSTA-AR-CCB-0 Watervliet, NY PERFORMING ORGANIZATION REPORT NUMBER ARCCB-TR SPONSORING / MONITORING AGENCY NAME(S) AND ADDRESS(ES) U.S. Army ARDEC Close Combat Armaments Center Picatinny Arsenal, NJ SPONSORING / MONITORING AGENCY REPORT NUMBER 11. SUPPLEMENTARY NOTES Presented at the SPIE 7 th Annual International Symposium on Smart Structures and Materials, San Diego, CA, March Published in proceedings of the symposium. 12a. DISTRIBUTION / AVAILABILITY STATEMENT Approved for public release; distribution unlimited. 12b. DISTRIBUTION CODE 13. ABSTRACT (Maximum 200 words) Modern tank guns, such as the one on the Abrams, are stabilized to allow fire on the move while traversing uneven terrain. The current barrel is short enough that treating it as a rigid beam allows engagement of another tank at ranges of over a kilometer. However, as the length of the tube is extended, to meet required muzzle exit velocities, the terrain-induced vibrations lead to increased muzzle pointing errors. A method to reduce these vibrations is to use the forward thermal shroud as part of a mass tuned damper. In this case the system under study is an extended length version of the gun currently fielded. This extended length increases its susceptibility to terrain-induced vibrations. The forward thermal shroud has been shortened and additional mass has been added onto its forward collar. This collar is then supported by springs, which are preloaded so that they stay in contact through the full range of the shroud's movement. Varying the stiffness of these springs allows for tuning of the absorber. Different types of springs and attachments have been tried. The current version uses leaf springs and a wedge collar. This system has been modeled, and experiments have been conducted to validate the model. 14. SUBJECT TERMS Vibrations, Vibration Absorber, Cannons, Gun Barrel, Passive, Accuracy, Dynamics 15. NUMBER OF PAGES PRICE CODE 17. SECURITY CLASSIFICATION OF REPORT UNCLASSIFIED 18. SECURITY CLASSIFICATION OF THIS PAGE UNCLASSIFIED 19. SECURITY CLASSIFICATION OF ABSTRACT UNCLASSIFIED 20. LIMITATION OF ABSTRACT UL NSN Standard Form 298 (Rev. 2-89) Prescribed by ANSI Std. Z

4 TABLE OF CONTENTS Page ACKNOWLEDGEMENTS iii INTRODUCTION 1 HISTORY 2 MATLAB MODEL 4 MODAL IMPACT TESTING 6 MATLAB MODEL WITH VIBRATION ABSORBER 8 VIBRATION ABSORBER TESTING 10 CONCLUSION 13 REFERENCES 14 TABLES 1. Frequency response parameters from modal analysis and the MATLAB" model 8 2. Peak location versus spring type Frequency response parameters from vibration absorber testing 12 LIST OF ILLUSTRATIONS 1. Original dynamically tuned shroud (DTS) approach 2 2. Second DTS concept 3 3. Current DTS concept 4 4. Barrel geometry 4 5. Test setup 5 6. Damped mode shapes and natural frequencies of the base system 6 7. Peak amplitude method 7 8. Predicted frequency response of the different absorbers 9

5 9. Relative frequency response of the different absorbers Frequency response of the different absorbers Relative frequency response of the different absorbers 11

6 ACKNOWLEDGEMENTS work: The authors acknowledge the following individuals for their help in conducting this Kenneth Olsen Paul Weber Franz-Peter Speckert Michael Gully Gary Cunningham David Smith. m

7 INTRODUCTION Modern tank guns, such as the one on the Abrams, are stabilized to allow fire while the vehicle is on the move, traversing uneven terrain. Vibration of the gun barrel leads to dispersion in the shot patterns. The wider the dispersion is, the more rounds are required to effect the desired damage on the enemy. An intuitive way to reduce this shot dispersion is to reduce the vibrations of the barrel. The end of the barrel is the anti-node for all vibration modes and its vibrations have the greatest effect on shot dispersion, so it is the obvious location to attempt to dampen the vibrations. This work focuses on doing just that. The system under study is the 120-mm M256E1 cannon. This is an extended length version of the M256 cannon currently fielded on the Ml Abrams. The M256E1 is 6.6 m in length compared to 5.3 m for the M256. This increase in length is necessary to achieve the required exit velocities. This increase in length, though, increases the barrel's receptance to environmentally induced vibration. (Receptance is the ratio of vibration amplitude to force as a function of frequency [ref 1].) While launch induced vibrations are certainly important, trying to structurally control the barrel during the available time window, typically less than eight milliseconds, is a daunting task beyond the scope of the present work. So, we will focus solely on environmentally induced vibrations. The absorber being considered is a spring collar that mounts onto the forward end of the gun's front thermal shroud. The primary function of the shroud is to prevent thermal gradients within the barrel that would otherwise cause unacceptable distortions (e.g., direct sunlight heating one side of the barrel). The back of the forward shroud is affixed to the bore evacuator using a compression collar that enables a pivoting action; this prevents thermal distortions in the shroud from being transmitted to the barrel. By connecting to the shroud, we are able to use the mass of the shroud as part of the mass of the absorber. The absorber is tuned by adjusting the mass and springs in the spring collar. By using a slightly shortened shroud, we are able to stay inside the space originally set aside for the thermal shroud and have minimal impact on system configuration. This approach was termed the dynamically tuned shroud (DTS) (ref 2). Originally, coil springs were used. Then these were replaced with leaf springs with adjustment screws and then leaf springs with a wedge collar. We will look at why these changes were made and how the current configuration was determined. Some of the problems with the current design will also be mentioned. The barrel is modeled in MATLAB using a finite element approach (ref 3). The Euler-Bernoulli finite element technique is used to generate second-order equations of motion of the barrel as a non-uniform beam. These are then converted to the first-order state space domain and transformed into the frequency domain. Predictions for the mode shapes and resonant frequencies are generated. After completing the model, it is verified by performing modal impact testing on the barrel. These results are then used to finetune the model.

8 The barrel is then tested with different vibration absorbers. Three different versions are used, the differences being the stiffness of the springs connecting the absorber to the barrel. These results will be compared to predictions of the MATLAB model. HISTORY The original concept for the DTS used compression springs mounted on the end of the thermal shroud. The spring collar itself had little added mass. To increase the mass, additional weights were added to the collar. To change the stiffness of the absorber, the preload of the springs could be changed. By altering these two parameters, the natural frequency of the DTS could be tuned. A picture of this approach can be seen in Figure l(ref 4). Figure 1. Original dynamically tuned shroud (DTS) approach. This approach was tested on a 120-mm XM291 cannon on the bump course at Aberdeen Proving Ground, Maryland, and it showed a 24 and 9% RMS improvement in vertical and horizontal bending, respectively (ref 5). No live fire testing was conducted, though, as there were doubts about the robustness of the hardware. The shroud being used had already been blown off one gun and the attachment of the spring collar was via ordinary screws. The effectiveness of this absorber spawned other efforts. A medium caliber absorber (ref 6) was designed and shown to be effective (ref 7). However, a more robust version of the large caliber DTS was desired. The goal was to come up with a weaponized version that could be tested in the field. During testing of the original DTS, it was found that adding weights to the end of the shroud increased performance. So, in the new design, the spring collar was to have a substantial mass, 15.8 kg. The profile of the compression springs was found to be high enough that they could interfere with the optics of the muzzle reference system (MRS). Also, there were questions about their reliability under use conditions. For these reasons, the second design used leaf springs. These springs were to be rigidly attached to the collar at one end and be supported by an adjusting screw at the other end. By turning this screw, the preload could be adjusted. To keep the entire assembly out of the line of sight

9 of the MRS optics, it was sized so that the absorber could move 3 mm in any direction. Three springs of different stiffness were selected so that they would remain in contact with the tube throughout this range of motion. All parts of the assembly were designed to withstand firing loads. A picture of this approach can be seen in Figure 2. Although four springs are visible in the picture, the collar was made to accommodate up to eight springs. Figure 2. Second DTS concept. This approach was supposed to be tested in-house and then sent to Aberdeen Proving Ground. The testing was to include both firing and non-firing tests on the bump course. However, problems arose during the in-house testing. The screws meant to preload the leaf springs did not ride smoothly on the surface of the tube and caused the spring to twist as it was loaded. Also, they tended to kick out and get stuck if the absorber was hit with enough mass to cause it to bottom. This type of loading would definitely be present during firing, so it was apparent that a redesign was needed. The third and current version of the DTS still uses the leaf springs but the adjusting screws are gone. Instead, a wedge ring is now mounted on the threads used for forward attachment of the standard thermal shroud. As it is screwed on, the wedge engages the springs and forces them out to a preset distance. This distance is set so that the absorber still has its full 3-mm range of motion in any direction. The springs are allowed to ride on the wedge ring, so there is no sticking or twisting as was found in the previous design. The apparent stiffness of the spring can be changed by turning the wedge ring in or out. The whole DTS still fits in the space of the normal thermal shroud and does not interfere with MRS optics. This is the design modeled and tested in the rest of the report. It can be seen in Figure 3. Originally, it was to be tested at Aberdeen in place of the second concept, but the additional time required to design and fabricate the new wedge ring prevented this from happening.

10 i i L i '' ' '' v -~' ; 1 *. ; rrrl - j! <L ~-~~ '- ^., Figure 3. Current DTS concept. MATLAB MODEL A finite element model of the barrel was created in MATLAB. Euler-Bernoulli beam approximations and Hermite-cubic interpolation functions are used to form the mass and stiffness matrices for the undamped second-order equations of motion by approximating the barrel, a continuous non-uniform beam, as a series of discrete elements. Continuity of lateral displacement and slope are imposed at the element boundaries. When assembled, these elements closely approximate the dynamics of the barrel (ref 4). The geometry of the barrel is entered in 1-mm increments and any non-circular cross sections are smeared together to become circular. The DTS hardware was added to the model as non-beam lumped masses. The mass of each part was distributed over its length, so that the center of gravity for each item was in the correct space. For this first version, the absorber was treated as another lumped mass with no springs. The mass of the beam is calculated by adding the mass of each of both the beam and non-beam slices. The model's version of tube can be seen in Figure 4. M25ÖE1 Profite & Non beam Masses Versus Length in Meters Total Non beam Mast of Kg, or bm c Linear Density from Profile Total Maas of Bum Kg. or 2SS7.0S4S bm O0 ' r-'x. ^- Figure 4. Barrel geometry.

11 After the geometry has been entered, the barrel is automatically broken into a user-defined number of elements. Nodes are forced to exist at both ends of the barrel and anyplace where constraints are specified. In this case, nodes were added at the location of the front and rear supports, the location where the hammer impulse would be delivered, and where the absorber will interact with the barrel. Even though the absorber wasn't in this version, this was done to keep changes between versions to a minimum. Ideally, we would like to hang the tube by soft springs. Originally, this was to be accomplished by striking the tube sideways with the supports acting as a pendulum. A sling suspended from a gantry was used near the breech end and multiple wraps of bungee cord suspended from a jib crane were used in the front. This setup can be seen in Figure 5. In this configuration, so long as the supports were long enough, very little energy would go into raising the barrel for small sideways displacements. Thus, the supports would approximate very soft springs. We were able to generate data for the non-absorber barrel in this configuration. Figure 5. Test setup. However, when it came time to test the barrel with the vibration absorber installed, we were unable to get any useful data. Testing in this configuration required the use of multiple springs on this absorber. We found that we were not able to introduce enough energy into the system to properly excite the absorber. To overcome this

12 problem, we changed to testing vertically, which required us to find the stiffness of the supports. This was found by suspending known weights from them and measuring the displacement. The effect of the less-than-soft rear support was minimized by placing it at the anti-node of the first mode. The location of the anti-node was found by running the model with no constraints. This location also turned out to be close to the anti-node for the second mode. Rayleigh proportional damping is used in the model. Initially, values from a previous report using this software for analyzing an XM291 gun barrel (ref 8) were used. After an experimental modal analysis on the barrel was conducted, experimentally found values were used and the model was rerun. Only minor differences in the resonant frequencies were found. After the required data were entered, the model was run and output generated. The software generates undamped and damped mode shapes and natural frequencies, a pole-zero plot of the eigenvalues, time response of the muzzle to a breech impulse, and a bode plot of the muzzle response, plus additional plots about the quality of the finite element analysis (FEA). In this case we are interested in the damped mode shapes and natural frequencies. These can be found in Figure 6. Rigid Body Mode Bend Mode Hz «Bend Mode 1, 1B.2873HI Bend Mode Hz Bend Mode Hz Axial Position (m) Figure 6. Damped mode shapes and natural frequencies of the base system. MODAL IMPACT TESTING After the MATLAB model was completed, an experimental modal analysis was done to validate the model. The barrel was hung from a sling at the breech and bungee cord at the muzzle. The idea was to simulate a free-free condition. As mentioned above, the rear support was placed at the anti-node of the first mode so as to limit its influence. As the supports were explicitly contained in the model, this was deemed satisfactory.

13 The goal of the modal analysis was to generate a frequency response plot between a force at the breech and the response of the muzzle. For this study an impact was used as the force and the acceleration of the muzzle was the response. An HP 3566A PC spectrum/network analyzer was used to calculate the frequency response. A PCB impact hammer with a super soft tip delivered the impact. The 6-dB roll off point of the tip was found to be 200 Hz. A PCB ICP accelerometer measured the response. The ICP power supply and signal conditioning for both of these were provided by a PCB 12- channel rack mounted power unit with a variable gain of 0 to 100 per channel. This can be seen in Figure 5. The HP 3566A was set up with a bandwidth of 800 Hz, 3200 frequency lines, and force/exponential windowing. Uniform averaging was performed, with a total of 10 averages being used per run. The gain was set to provide good signal strength. After each impact, the data were checked for double hits and overloading of the accelerometer. The peak amplitude method (ref 1) was used to extract the necessary modal parameters from these data. To determine the damping ratio of a peak, equations (1) and (2) were used: <;=y 2 n 77 = ^ col-col 2 cot (1) (2) Where "is the viscous damping ratio, r\ the structural damping loss factor, (Or is the natural frequency of the peak, and ü) a and C0b are the half power points. These quantities can be seen in Figure 7. Figure 7. Peak amplitude method. Once ^has been found for at least two peaks, the proportional damping coefficients, a and ß can be found from the following formulas:

14 a = -2-0)\ + co\^\ - Q)\Q + Co\ (3) -a+acl + l^^c: 20) 2 co 2 Using these formulas, we extracted the data in Table 1 from the frequency response data. Table 1. Frequency response parameters from modal analysis and the MATLAB model Peak Magnitude Frequency Coherence c MATLAB 10 db (g/lbf) Hz Hz (4) a(s-') ß(s) 1.899xl0" b Comparing the experimental values with those from the MATLAB model, we see that there is good agreement between the two. Given that the resolution of the analyzer is 0.25 Hz, we can say that the first mode is found exactly. The other modes are off by about 1 Hz. In the higher modes the system is found to actually be slightly stiffer than predicted. This is possibly attributable to the less than ideal boundary conditions. The calculated a and ß are used in all the future models. MATLAB MODEL WITH VIBRATION ABSORBER Now that the model has been validated for the plain barrel, it must be modified to include the vibration absorber. As mentioned earlier, the absorber is a proof mass type actuator that connects to the forward thermal shroud. The assembly consists of a shortened forward thermal shroud, onto which mounts the spring collar, onto which mount the leaf springs, which are in turn supported by the wedge collar. The shroud is m in length and has a mass of 21 kg. The collar has a mass of 15.8 kg and has spaces for mounting up to eight springs. The leaf springs are trapezoidal in shape and 86- mm in length. The wedge collar screws onto the barrel threads for the standard thermal shroud. Its purpose is to preload the springs evenly. The springs should never lose contact with the wedge ring during normal operation. Though up to eight springs can be used at once, only one was used in the model and during testing. This spring was mounted in the top location. Springs of three different spring constants were used in both the model and experiment. The three different spring types had thicknesses of 2-, 2.8-, and 3-mm, and spring constants, as

15 ordered, of 87563, , and N/m, respectively. However, these spring constants were for the springs under conditions of normal leaf spring use. With our setup this is not the case, so we installed each spring type, applied known weights, and measured the deflection to get an apparent spring constant for our model. The measured spring constants for the 2-, 2.8-, and 3-mm springs were , , and N/m, respectively. The shroud was treated as a spring and, thus, by the standard approximation for a spring with mass, 1/3 of its mass was added to the absorber mass and 2/3 was added to the barrel as a lumped mass. The location of the lumped mass was adjusted so the center of gravity of the shroud and absorber mass together were positioned as in the actual assembly. The MATLAB model allows for a mass and stiffness to be entered for a vibration absorber. The mass was a combination of the absorber mass and 1/3 of the shroud mass. These masses were removed from the non-beam mass part of the model, so they wouldn't be counted twice. A separate model was made for each spring type. The stiffnesses found above were entered for the corresponding model. The resulting natural frequencies for each spring type were 13.92, 17.27, and Hz for the 2-, 2.8-, and 3- mm models. As with the plain barrel, the models were run once all required data were entered. Damped mode shapes and natural frequencies were recovered, along with bode plots and pole-zero maps. Figure 8 shows the predicted frequency responses of the different absorbers in terms of the inertance. (Inertance is the ratio of vibration acceleration to force as a function of frequency [ref 1].) Figure 9 shows the relative response of each absorber as compared to the original configuration. Whenever the trace is below the 1.0 line, the response is less than the base system. If it is above the 1.0 line, then the response is greater. M256E1 Frequency Response for Different Absorber* 1T from Breech Impact - Muzzle Response Figure 8. Predicted frequency response of the different absorbers.

16 ' ' i.e 1.4 ' M256E1 RaWva. RasponM ORarent Absorbars v /. /',' S^-- """-- s t COS i ; / - 04 ;j ;\ ' i \ 02 I Fnsquemcy (Hi) 1 Nor* mm Spring mm Spring j mm Spring Figure 9. Relative frequency response of the different absorbers. From Figure 8 the effects of the different absorbers on the first mode are readily apparent. The absorber is removing the first peak of the barrel and shifting its energy into new peaks on either side of it, as expected (ref 9). The two plots also show that the absorber only affects the peak nearest it. The second and later peaks are not affected to any real degree. This is really noticeable in the relative response where the responses quickly return to a value of one after the first mode. At present we are not sure why the trace for the 2.8-mm spring does not follow this trend. As it appears to be evenly offset from the rest of the traces, we feel it is a coding error, but have not been able to find it. Table 2 summarizes the locations of the modes for each of the absorbers. Peak A is the new peak created by the absorber. Table 2. Peak location versus spring type Peak None 2.0-mm 2.8-mm 3.0-mm Hz Hz Hz Hz A VIBRATION ABSORBER TESTING Now that we have a model that includes the vibration absorber, modal analyses were done to validate it. The absorber was tested with each of the different spring types. The barrel orientation and accelerometer placement were kept the same as for the last plain barrel test. This ensured that any changes in the frequency response should be 10

17 directly attributable to the vibration absorber and not changes in test setup. The same procedure mentioned previously was followed for the tests. The recovered frequency responses in terms of inertance are shown in Figure 10. Figure 11 shows the relative response of the absorbers as compared to the non-absorber system. Comparing these figures back to Figures 8 and 9 makes some things immediately apparent. First the expected notch for the absorber is missing from the 2.0- and 2.8-mm configurations. There is a slight notch on the 3.0-mm version. Graphically, the shifting of the barrel's first peak and the lack of influence on the higher modes is still apparent though. M256E1 Experimental Frequency Response 12" from Breech Impact Muzzle Response None mm Spring 2.8 mm Spring ~ mm Spring s i J\ \ i / \t ft. /' ' \v / V <A A."J *\ ^***^r ^<^Z'^. * 1 /v-~u r ww ', // y \ Vi. s ' V' Figure 10. Frequency response of the different absorbers., ft **\- 1 ' N- :: '' -...;; fc " ::: V / 1 ',' V/ ; / '1 i i None mm Spring 2.6 mm Spring mm Spring Frequency (Hz) Figure 11. Relative frequency response of the different absorbers. Table 3 lists the frequency response parameters recovered from the experimental data. The data for the non-absorber configuration are repeated here to make comparisons easier. The same observations that can be made from comparing the graphs are just as 11

18 apparent by comparing Tables 2 and 3. Let's look at each of the configurations individually. Table 3. Frequency response parameters from vibration absorber testing No Vibration Absorber Peak Magnitude Frequency Coherence db (g/lbf) Hz 1 ^ mm Vibration Absorber Peak Magnitude Frequency Coherence db (g/lbf) Hz 0 1 ^ mm Vibration Absorber Peak Magnitude Frequency Coherence db (g/lbf) Hz 0 1 ^ mm Vibration Absorber Peak Magnitude Frequency Coherence db (g/lbf) Hz ^ The 2.0-mm absorber is missing the expected notch. From the MATLAB model, the notch should be at Hz; however, Figure 10 shows nothing at that frequency. The location of the next peak, though, matches between model and experiment. Both 12

19 show it being shifted from 19.3 to 20.3 Hz. They also show the absorber's effects being only noticeable around the first mode of the non-absorber barrel. The 2.8-mm absorber shows the same trends as the 2.0-mm absorber. It, too, is missing the absorber notch and lower peak; this time it should be at Hz. Peak 1, though, once again matches between the model and experiment. Also, the effects of the absorber are again localized around the affected mode, as they should be. Neither the 2.0- nor the 2.8-mm absorbers, though, appear to have any effect on the magnitude of the response around barrel's first mode, as would be expected. The 3.0-mm absorber tells a slightly different story. This time the notch is there. It is small, but it is detectable. The lower peak is at Hz, which doesn't match with the model but is the calculated natural frequency of the absorber. The next peak does match with expectations. Over the frequency range between these two peaks, the response is reduced by about 4 db, which is noticeable though smaller than one would desire. Like in the other two configurations, the absorber's effects drop off quickly as one moves away from affected mode. The lack of a notch for the absorbers is puzzling. They are having an effect on the barrel's mode near their natural frequencies by pushing it to a higher frequency, but only the 3.0-mm one is having a noticeable effect on the magnitude. We are not sure what is causing this behavior. One idea we have is that we are not getting enough energy into the system through the impact hammer. Without enough energy, the friction and other nonlinearities in the components coupling the DTS to the system may not be overcome and, thus, it would not function properly. If this were correct, this would be a lab only type problem, as getting energy into the system would not be a problem in the test vehicle. CONCLUSION This report has explored the effect of mounting a vibration absorber to the forward thermal shroud of a 120-mm M256E1 cannon. AMATLAB model of the nonabsorber barrel was developed and then verified by performing modal impact testing upon the actual barrel with the absorber rigidly affixed. Good agreement was found between the model and experimental data. After modeling and testing the non-absorber barrel, a vibration absorber was modeled and tested to find its effects upon the barrel's frequency response. Three different configurations were modeled and tested. Decent agreement on the location of the modes was found between model and experiment. However, the notch type effect of the absorber was missing from two of the three configurations. In the third it was noticeable but small. Only this third version attenuated the system response to any degree. No explanation for this lack of absorber notch is readily apparent and verifiable. One possible one is that not enough energy is getting into the system through the impact hammer to overcome the friction and other nonlinearities in the components coupling the DTS to the system. At present, though, we have no plans on testing this theory, as there is no funding for more testing. The original version of the DTS, though, was tested on a 13

20 vehicle and had no problems with getting enough energy into the system. If future testing of the DTS or similar devices is undertaken, steps will be taken to make sure more energy can be gotten into the system. If done in the lab then most likely we will order a shaker or a change in boundary constraints. For testing in a vehicle, getting energy into the system has never been a problem. REFERENCES 1. Ewins, DJ., Modal Testing: Theory, Practice and Application, Research Studies Press, Ltd., Baldock, England, Käthe, E., "Gun Barrel Vibration Absorber," U.S. Patent , January Kathe, E., "MATLAB Modeling of Non-Uniform Beams Using the Finite Element Method for Dynamics Design and Analysis," U.S. Army ARDEC Report ARCCB-TR-96010, Benet Laboratories, Watervliet, NY, April Kathe, E., "Lessons Learned on the Application of Vibration Absorbers for Enhanced Cannon Stabilization," Shock and Vibration, Vol. 8, 2001, pp Kathe, E.L., "Performance Assessment of a Synergistic Gun Barrel Vibration Absorber During Bump-Course Testing," ARDEC Technical Report ARCCB-TR , Benet Laboratories, Watervliet, NY, September Kathe, E., "Muzzle Break Vibration Absorber," U.S. Patent Application Serial Number 09/ , filed 5 July Littlefield, A., Kathe, E., Messier, R., and Olsen, K., "Gun Barrel Vibration Absorber to Increase Accuracy," Proceedings of the 42" AIAA/ASME/ ASCE/AHS/ASC Structures, Structural Dynamics, and Materials Conference, Seattle, WA, Apr Kathe, E., "Design and Validation of a Gun Barrel Vibration Absorber," Proceedings of the 67th Shock and Vibration Symposium, Vol. 1, Published by SAVIAC, Monterey, CA, November 1996, pp den Hartog, J.P., Mechanical Vibrations, McGraw-Hill Book Company, Inc., New York,

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